FIG. 6 is a conceptual view of the configuration of a generally used conventional speed control device for a motor. In the figure, reference numeral 1 denotes a motor, reference numeral 2 denotes a load driven by the motor, and reference numeral 3 denotes an axle connecting the motor 1 and the load 2. The load 2 represents a moveable member of a machine driven by the motor 1, modeled as an inertial load, and the axle 3 models a mechanism for transmitting torque, generated by the motor 1, to the machine. Reference numeral 4 denotes an encoder, attached to the motor 1, for detecting the position of the motor 1, reference numeral 5 denotes a speed detection means for computing motor speed by differentiating the motor position detected by the encoder 4. Reference numeral 6 is a comparator for comparing a speed command signal given by a high-level controller (not illustrated in the figure) with the motor speed output by the speed detection means 5, and for outputting the difference between these two speeds as a speed error. Reference numeral 7 is a speed control means that, with the speed error output by the comparator 6 as input, outputs an electrical current command that is a motor drive command. Reference numeral 8 is an electrical current control means that, by controlling the motor electrical current based on the electrical current command output by the speed control means 7, generates torque in the motor and rotates the motor.
The speed control means 7 is composed of a proportional controller 9, an integral controller 10, and an adder 11. The proportional controller 9 multiplies the input speed error by proportional gain KP and outputs the result; and the integral controller 10 multiplies the integral value of the speed error by integral gain KI and outputs the result. The adder 11 adds the proportional controller 9 output and the integral controller 10 output, and outputs the result as the electrical current command. Reference numeral 15 is a mechanical system and is comprised of the motor 1, the load 2, and the axle 3.
A conventional speed control system is configured as described above; the torque for accelerating or decelerating the motor 1 is generated so that the speed error, being the difference between the speed command signal and the motor speed, becomes small, and thus the motor 1 and the load 2 rotate so that the motor speed follows the speed command signal given by the high-level controller. In cases where disturbing torque acts in the mechanical system, the motor speed fluctuates due to the disturbing torque; however, this speed fluctuation is detected by the encoder 4 and the speed detection means 5, is fed back to the speed control means 7, and an electrical current command is generated to correct the speed fluctuation. In this way, even in cases where disturbing torque acts, the speed fluctuations are restrained by a speed control loop, and the motor is controlled so as to follow the speed command signal.
FIG. 7 illustrates frequency response for an open speed loop in a configuration for a speed control system using a conventional speed control device. Frequency response for a speed open-loop are frequency response from the input of the speed control means 7 to the motor speed detected by the speed detection means 5; the upper graph in FIG. 7 illustrates gain characteristics and the lower graph illustrates phase characteristics. The broken lines in the figure illustrate frequency response for cases where the rigidity of the axle 3 connecting the motor 1 and the load 2 is high, that is, the mechanical rigidity is high. The full lines illustrate frequency response for cases of mechanical resonance, where the mechanical rigidity is low because the rigidity of the axle 3 is low.
Gain characteristics for cases where the mechanical rigidity is high (the rigidity of the axle 3 is high), as illustrated by the broken line in the upper graph of FIG. 7, drop down towards the right, over all frequencies. As illustrated by the broken line in the lower diagram of FIG. 7, at high frequencies, phase lags become large due to phase lags resulting from the sample period of the electrical current control means 8 or the control device; and at low frequencies, phase lags also become large due to using the integral controller 10 in the speed control device 7.
On the other hand, in cases where the mechanical rigidity is low (the rigidity of the axle 3 is low), the mechanical system has a mechanical resonance, and, as illustrated by the full line in the upper diagram of FIG. 7, its gain characteristic has a resonance referred to as a peak and an anti-resonance referred to as a trough. Since the gain decreases to the left, between resonance and anti-resonance, the gain at low frequencies becomes small in comparison to cases where the rigidity of the axle 3 is high (the broken line). As the inertial moment J of the whole mechanical system, being the sum of the inertial moment of the motor 1 and the inertial moment of the load 2, becomes large in comparison to the inertial moment JM of the motor, the distance between the resonance and the anti-resonance increases and the frequency of the anti-resonance becomes low, so that the low frequency gain becomes even smaller.
In order to realize high accuracy speed control, it is necessary to curb the influence of the disturbing torque that is one of the causes of the speed error, and to make small the speed fluctuations when the disturbing torque acts. In general, since the disturbing torque has low frequency signal components, in order to make small the speed fluctuations due to the disturbing torque, it is necessary to enlarge the gain at low frequencies. However, with machines having low rigidity, since the gain is small at low frequencies as described above, the speed fluctuations due to the disturbing torque become large and high accuracy control becomes difficult.
To enlarge the gain at low frequencies in the conventional speed control devices illustrated in FIG. 6, it is necessary to enlarge the proportional gain KP and the integral gain KI of the speed control means 7. However, there is a limit to this, and ultimately high accuracy control could not be realized. The reason for this is explained below. In FIG. 7, the frequency, where the gain characteristic for low rigidity cases (full line) intersects the line for 0 db gain at frequencies lower than the anti-resonance frequency, is a first crossover frequency ωC1, and the frequency where it intersects the line for 0 db gain at frequencies higher than the resonance frequency is a second crossover frequency ωC2. In order that the control system operates stably without causing vibrations or oscillating, it is necessary that the speed open-loop frequency characteristic phase lags be small for the first crossover frequency ωC1 and the second crossover frequency ωC2. However, when the proportional gain KP is made large, since the second crossover frequency ωC2 shifts to the high frequency side, the phase lag for the second crossover frequency ωC2 becomes large and the control system ends up vibrating and oscillating. When the integral gain KI is made large, since the phase lag at low frequencies due to the integral controller 10 becomes large, the phase lag for the first crossover frequency ωC1 becomes large, and the control system ends up vibrating and oscillating. In this way, since for the conventional speed control device it is not possible to enlarge the proportional gain KP or the integral gain KI beyond a certain amount, the gain at low frequencies cannot be made large, and as a result high accuracy control has been difficult.
Servo control technology directed towards mechanical systems having mechanical resonance is disclosed in Japanese Laid-Open Patent Publication 2000-322105. FIG. 8 illustrates a speed control device using this technology. The same reference numerals are used for members that are the same as in FIG. 6. In FIG. 8, reference numeral 12 is a filter inserted serially in the speed control loop and has a reverse characteristic or a proximately reverse characteristic to the anti-resonance/resonance characteristic of the mechanical system. This filter is adjusted to exhibit a characteristic such as that described in equation (1).G(s)=ω12(s2+2ζ2ω2s+ω22)/{ω22(s2+2ζ1ω1s+ω12)}  (1)Here, ω1, ω2, ζ1, and ζ2 each are parameters, ω1 has a value close to the anti-resonance frequency, and ω2 has a value close to the resonance frequency. ζ1 and ζ2 are arranged to be small, according to the anti-resonance and resonance peaks. According to this technology, since it is possible to restrain the gain of the resonance peak of the mechanical system by the filter 12, the gain can be raised more than in conventional cases, and high accuracy control can be realized.
However, this technology is mainly directed less at the resonance peak illustrated in FIG. 7, than at restraining the unstable resonance peak at higher frequencies, and it is inexpedient to apply this technology to stable resonance peaks such as in FIG. 7. The reason for this inexpediency is explained below.
FIG. 9 illustrates the frequency response of the filter 12 in cases where this technology is applied to a mechanical system having frequency response illustrated by the full lines in FIG. 7. Since the filter 12 is adjusted to have characteristics, the reverse of the anti-resonance/resonance characteristics of the mechanical system as described above, the gain characteristic has a peak at the anti-resonance of the mechanical system and has a trough at the resonance of the mechanical system. The case where the disturbing torque acts on the mechanical system controlled using this type of filter will be looked at. The resonance frequency is a frequency at which the mechanical system vibrates easily, and when the disturbing torque acts, the mechanical system may vibrate at the resonance frequency. This vibration is detected by the speed detection means 5 and is fed back to the speed control means 7; the speed control means 7 generates and outputs an electrical current command in order to stop this vibration. A signal component at the resonance frequency is, of course, included in this electrical current command. However, since the gain of the filter 12 becomes small at the resonance frequency as illustrated in FIG. 9, by passing through the filer 12, the signal component at the resonance frequency ends up being removed from the electrical current command. That is, the signal component for stopping the vibration ends up being removed, by the filter 12 from the electrical current. As a result, the disadvantage occurs that even when the mechanical system vibrates at the resonance frequency, the speed control system cannot stop this vibration.
As described above, when controlling a mechanical system of low rigidity and where the inertial moment J of the mechanical system is large in comparison to the inertial moment JM of the motor, there have been problems in that it is difficult to make the gain large at low frequencies with the conventional generally-used speed control device, and speed fluctuations due to disturbing torque become large so that high accuracy control is difficult.
Furthermore, the technology disclosed in Japanese Laid-Open Patent Publication 2000-322105 enables the gain to be made large at low frequencies but cannot stop mechanical vibration, so that it could not be applied to mechanical systems having mechanical vibration at low frequencies.